Both mechanical spring suspensions and air spring suspension systems are commonplace in heavy duty vehicles. Heretofore, both types of suspension systems have neglected for many purposes the results of applying a very large load to the suspension system. For example, truck tractors are often used to pull large semi-trailers with heavy loads. The semi-trailer may place on the truck suspension system loads that are ten times the load exerted thereon by the unloaded truck tractor.
To consider the effects of such loads, one may consider simplified dynamic models.
Consider first a mechanical spring (e.g., coil or leaf spring) system comprising a mass M1, a spring having stiffness K lbf/in, and a damper having damping coefficient C lbf/in/sec. As an example, suppose the system values are: ##EQU1##
These are fairly typical values for one "corner" of a tractor leaf spring rear suspension.
Then the system has an undamped natural frequency of ##EQU2## critical damping of ##EQU3## and a damping ratio of ##EQU4##
Now, suppose that a large mass is added to the system, such that the total mass M2 is substantially larger, say, ten times larger, than original mass M1. This is in fact the order of change in the load supported by the rear suspension that occurs when hooking up a laden semi-trailer to a tractor. Under the additional load, the spring deflects by an amount ##EQU5## Now the undamped natural frequency is ##EQU6## the critical damping is ##EQU7## and the damping ratio is ##EQU8##
Three things are immediately apparant:
(1) A high degree of spring stiffness is essential to avoid excessive static deflection under load.
(2) The combination of low mass and high stiffness results in a very high natural frequency in the light load condition, meaning that this system will be rather ineffective as a vibration isolator in the light load condition.
(3) A damping ratio that was optimal for the light load condition, gives only about 30% of the desired damping under the heavy load condition.
As a next model, consider a system utilizing an air spring and active leveling. In some manner, all such air suspension systems interpose an air spring between the vehicle axle and the frame. In a typical such system, suspension brackets are attached to the vehicle frame, and the lower ends of the brackets are pivotally connected to link assemblies made up of a flexible mechanical spring member and a rigid paddle member. The link assemblies are typically attached to the vehicle axles by U-bolts, while air springs of the rolling lobe type are interposed between the paddle member and the vehicle frame.
Most air spring suspension systems also employ active leveling in which the truck's high-pressure air supply is connected to the input of a leveling valve which, by means of a mechanical linkage, senses the distance between the vehicle frame and the axle. When the frame is low with respect to the axle, then, after a suitable delay, the leveling valve opens and admits air to the air springs, raising their internal pressure and causing the frame to rise until the desired ride height is reached. Similarly, if the frame is too high, relative to the axle, the leveling valve exhausts air from the air springs, reducing their internal pressure and causing the frame to lower to the desired ride height. It is usual to have a single leveling valve supply all air springs, so that the pressure in all the springs is the same.
Now consider such an air spring system employing active leveling. Suppose that the system values are: ##EQU9##
Again, these are fairly typical values for one "corner" of a tractor rear air suspension when unladen.
The system has an undamped natural frequency of ##EQU10## critical damping of ##EQU11## and a damping ratio of ##EQU12##
If, again, the mass of the system is increased by a factor of ten, there is extreme settling due to the low spring rate. However, the leveling valve senses this, and increases the air pressure in the air spring until the desired height is reached. Since air springs of the type under discussion maintain fairly constant bearing area regardless of their pressure, it follows that, since ten times more mass is supported in the fully loaded condition than in the light condition, the air spring pressure will also be ten times greater than the original pressure. This is a very important point: in air spring systems, the spring air pressure is proportional to the load supported.
It is inherent in the physics and thermodynamics of air springs of the type under discussion that the spring rate is directly proportional to the spring air pressure, and thus in turn is directly proportional to the mass supported. In general, if in supporting mass M1 there is a spring rate K1, in supporting a mass M2 there will be a spring rate of EQU K.sub.2 =M.sub.2 /M.sub.1 K.sub.1
In the example just considered, therefore, in the heavy load condition the spring rate will be K2=1850 lbf/in. The undamped natural frequency will be ##EQU13## critical damping will be ##EQU14## and the damping ratio will be ##EQU15##
Since the spring air pressure increases with the mass supported, and since spring stiffness increases with spring pressure, the natural frequency, unlike the mechanical spring system, is low even when unladen, and remains constant in going from the unladen to the laden condition. These are very desirable features from a vibration isolation standpoint, and are the principal reasons why air spring suspensions have gained such wide acceptance in the truck field.
However, in going from mass M1 to mass M2, the critical damping for the mechanical spring system is increased by the square root of the ratio of the masses, i.e., ##EQU16## for a mechanical spring system, while for the air spring system, critical damping increases directly with the ratio of the masses, i.e., ##EQU17## for an air spring system. For this reason, a mechanical spring system which has its mass increased by a factor of ten will still retain about 30% of the desired damping level, while an air spring system subject to the same mass increase will retain only about 10% of the desired damping level.
When suspension systems have constant damping coefficients and support varying quantities of mass, they suffer some degradation in performance, but air spring systems suffer more degradation than mechanical spring systems.
Heretofore in this discussion it has been assumed that the damping is linear, i.e., that the dampers provide a force linearly proportional to the relative velocity of the axle and frame. In fact, automotive dampers are intentionally made non-linear.
As long as the relative velocity of the frame and the axle is small, as in normal minor oscillations of the suspension, a linear damper can be made to work well, but the suspension must occasionally traverse large obstacles such as rocks or lumber in the roadway. Under such circumstances, which involve very high relative velocities of the frame and axle, a linear damper tends to transmit excessive, even destructive force to the vehicle frame. Therefore most automotive dampers employ some sort of "blow-off" valving such that above a certain piston speed the damper force increases only slightly with velocity.
Most automotive dampers are non-linear in yet another way: since the shock loads caused by striking obstacles are generally associated with upward motions of the axle, the damping coefficient for jounce (compressing) travel is generally made less than the damping coefficient for rebound (extending) travel, the theory being that the light jounce damping allows the wheel to "get out of the way" of the obstacle, while the heavier rebound damping ensures that any subsequent oscillations of the wheel will be quickly damped out.
A linear damper with equal jounce and rebound damping has elliptical dynamometer curves, whereas a velocity-softening damper has more nearly rectangular dynamometer curves.
A general object of the invention is to provide improved ride characteristics for vehicles having wide variations in load.
An important object of the present invention is to provide a damper, for both mechanical spring suspensions and for air spring suspension systems, such that the damper has a damping coefficient which is variable and is always proportional to the quantity of mass supported. This is accomplished in air-suspension systems by using the suspension system air pressure, which is always proportional to the mass supported, to modulate the damping. For mechanical systems a special hydraulic pressure is generated, is made proportional to the mass supported, and is applied to modulate the damping.
A further object of the invention is to accomplish the preceding object while retaining the other features commonly found in automotive dampers, i.e., high speed blow-off valving, and different damping coefficients in the jounce and rebound directions.